1. Field of the Invention
This invention relates to an oil-flooded screw compressor of the type which includes a balance piston for causing a force to act upon a screw rotor in a direction from the suction side to the discharge side, and a slide valve for adjusting the volume of the screw compressor.
2. Description of the Prior Art
Screw rotors of a screw compressor, particularly a male rotor, is acted upon, during compressing operation of the screw compressor, by a great thrust force in a direction from the discharge side to the suction side due to the structure of the screw compressor. Where such thrust force is excessively great, it will significantly reduce the life of thrust bearings by which the screw rotors are supported for rotation.
An oil-flooded screw compressor wherein a thrust force acting on a thrust bearing is reduced has been proposed and is disclosed, for example, in Japanese Utility Model Laid-Open No. 175592/1986. The thus disclosed oil-flooded screw compressor is shown in FIG. 6.
Referring to FIG. 6, the screw compressor includes a pair of female and male screw rotors 5 accommodated in a casing 3 and supported for rotation by means of a pair of bearings 4a and 4b. The casing 3 has a suction port 1 formed at an end thereof and has a discharge port 2 formed at the other end thereof. A balance piston 7 is provided at an end of a suction side rotor shaft 6 of one of the screw rotors 5 and fitted for sliding movement in a cylinder chamber 8 formed in the casing 3.
An oil separating and collecting device 32 is interposed in a discharging flow path 31 connecting to the discharging port 2, and an oil flow path 36 extends from an oil storage portion 33 at the bottom of the oil separating and collecting device 32. An oil cooler 34 and an oil pump 35 are interposed in the oil flow path 36, and the oil flow path 36 is branched into two paths and are communicated, on one hand, with lubricating portions of shaft seal parts, the bearings 4a and 4b and so forth by way of a flow path not shown and, on the other hand, with the cylinder chamber 8 by way of a pressure oil supplying port 9.
With the screw compressor, gas sucked into a gas compressing spacing in the inside of the casing 1 by way of the suction port 1 is compressed by the screw rotors 5 and discharged by way of the discharging port 2 together with oil for the cooling and so forth which has been inadvertently admitted into the gas compressing spacing. Then, the gas and oil thus discharged are introduced into the oil separating and collecting device 32 in which they are separated from each other. The compression gas from which the oil has been removed is sent out from an upper portion of the oil separating and collecting device 32. On the other hand, the oil drops into and is stored in the oil storage portion 33. Then, the oil is sent out from the oil storage portion 33 and then cooled by the oil cooler 34, whereafter it is fed to the lubricating portions and the end of the cylinder chamber 8 remote from the screw rotors 5. The oil admitted into the spacing around the screw rotors 5 is thereafter circulated along a similar route so that it may be used after then.
As oil of the oil flow path 36 is introduced to the end of the cylinder chamber 8 remote from the screw rotors 5 in this manner, a thrust force acting upon the screw rotors 5 from the discharge side to the suction side during operation of the screw compressor is reduced so that an excessive force may not be applied to the bearing 4b.
With the conventional oil-flooded screw compressor, the oil pressure at the end of the balance piston 7 remote from the screw rotors 5 is substantially equal to a discharge pressure Pd at the discharge port 2. However, if the other end of the balance piston 7 adjacent the screw rotors 5 is communicated directly with the suction port 1, then gas containing oil therein will flow from the cylinder chamber 8 to the sucking port 1, whereupon it is expanded, which will result in reduction of the amount of gas to be sucked into the rotor chamber by way of the sucking flow path. Therefore, the sucking port 1 is communicated with a gas enclosing spacing having an inner pressure a little higher than a suction pressure Ps which will appear where it is not communicated with the gas enclosing spacing, for example, a gas enclosing spacing having a pressure of 1.3Ps. Accordingly, a force F acting upon the balance piston 7 in a direction from the suction side to the discharge side is represented by the following expression, and during operation of the screw compressor, the magnitude of the force F is fixed when the discharge pressure Pd and the suction pressure Ps are fixed. EQU F=S.multidot.(Pd-1.3 Ps)
where S represents an area of the pressure receiving portion of the balance piston 7. Here, a sectional area of the rotor side shaft 6 is ignored.
By the way, where the screw compressor is of the type which has a volume adjusting slide valve, the thrust force produced at the screw rotor 5 is reduced during partial load operation or no load operation of the screw compressor comparing with that during full load operation, and a force acting upon the balance piston 7 due to the oil pressure and another force acting upon the rotor shaft 6 from the screw rotors 5 sometimes become substantially equal to each other, which may put the bearing 4b into a condition wherein it undergoes so little thrust load that it may drift.
Referring to FIG. 7, the axis of abscissa indicates a slide valve position in a ratio (%) of the load in an operating condition at the position to the full load while the axis of ordinate indicates a force acting upon a thrust bearing. When the force acting upon the bearing 4b becomes excessively great until it exceeds a predetermined value f.sub.1, the life of the bearing becomes shorter than a fixed reference interval of time, for example, 20,000 hours. Thus, while the force where the balance piston 7 is not provided is such as shown by an alternate long and two short dashes line curve I which exceeds the force f.sub.1 when the slide valve comes to a position considerably near to its full load position (100%), according to the screw compressor described above in which the balance piston 7 is provided, the force acting upon the bearing 4b is reduced uniformly by a same magnitude over every position of the slide valve such that the highest value thereof may be smaller than the level f.sub.1 as seen from another solid line curve II in FIG. 7.
However, if the force acting upon the bearing 4b is excessively small below another predetermined value f.sub.2, then the bearing 4b may drift and be likely damaged. In particular, even if such balance piston 7 as described above is provided, a problem still remains that, if the slide valve approaches the no load operation position (0%) as seen from the curve II, the force acting upon the bearing 4b becomes smaller than the value f.sub.2 and is liable to be damaged.